Pumping cavity for rotary vane pump



Aug. 16, 1960 D. J. DESCHAMPS 2,949,081

PUMPING CAVITY FOR ROTARY VANE PUMP Filed April 25, 1956 7 Sheets-Sheet 1 FIG! FIG. 2

.- i I IN VEN TOR.

DESIRE J. DESCHAMPS ATTORNEY Aug. 16, 1960 D. J. DESCHAMPS PUMPING CAVITY FOR ROTARY VANE PUMP Filed April 25. 1956 7 Sheets-Sheet 2 INVENTOR.

DESIRE J. DESCHAMPS ATTORNEY Aug. 16, 1960 D. J. DESCHAMPS 2,949,081

PUMPING CAVITY FOR ROTARY VANE PUMP Filed April 25, 1956 7 Sheets-Sheet 3 FIG. 4

INVENTOR.

DESIRE J. DESCHAMPS BY AT TORNEY Aug. 16, 1960 D. J. DESCHAMPS PUMPING CAVITY FOR ROTARY VANE PUMP Filed April 25, 1956 '7 Sheets-Sheet 4 INVENTOR.

DESIRE J. DESCHAMPS ATTORNEY Aug. 16, 1960 D. J. DESCHAMPS 2,949,031

PUMPING CAVITY FOR ROTARY VANE PUMP Filed April 25, 1956 '7 Sheets-Sheet 6 OUTLET L I FIG. 12 g i l FIG.13 i

INVEN TOR.

DESIRE J. DESCHAMPS ATTORNEY Aug. 16, 1960 D. J. DESCHAMPS PUMPING CAVITY FOR ROTARY VANE PUMP Filed April 25, 1956 7 Sheets-Sheet 7 IN VEN TOR.

DESI RE J. DESCHAM PS ATTORNEY v Un ted States Pa ent f 2,949,081 PUMPING CAVITY FOR ROTARY VANE PUMP Desire J. Deschamps, San Fernando, Calif., assignor to Hydro-Aire, Inc.,' Burbank, Calif., a corporation of California Filed Apr. 25, 19 55, ser. No. 580,482 5 7 Claims. cams-120 This invention relates to rotary .vane pumps and, more particularly, to an improved design of cavity liner or cylinder which, .although of greatest advantage in both fixed and variable capacity high pressure pumps, is also useful in low and medium pressure pumps;

Rotaryvane pumps, as will be understood. by those skilled in the. art, comprise a substantially cylindrical rotor mounted in a larger diameter, embracing or encompassing cylindrically surfaced cavity. The axes of the rotor and cavity cylindrical surfaces are offset or eccentric relative to each other, so that parts of the rotor circumference are closer to the cavity surface than are other parts of the rotor circumference. Substantially flat vanes are slidably mounted in the rotor with their outer edges engaging the surface of the cavity, these vanes general-1y being mounted for movement in radial or axial planes of the rotor.

As the rotor rotates, the vanes, due to the varying distance between the cylindrical surfaces of the rotor and of the cavity, are moved between extended and retracted positions so that the volume enclosed between circumferentially adjacent vanes is substantially continually varied in amount. The parts are so arranged that vanes travelling toward the outlet port are retracted to decrease the volume of the fluid between adjacent vanes and thus force it under pressure through the outlet port. Similarly, in travelling toward the inlet port, the .vanes are projected to increase the volume between adjacent vanes.

Various means have been used to maintain the vane outer edges in engagement with the cavity surface. .For example, some pumps employ vanes extending in an axial plane completely'through the rotor so that both outer edges of the vanes are engaged at all times with the cavity surface which, in this case, is slightlyelliptical in shape. Other pumps use vanes mounted in radial slots so as to project only from one side of the rotor and, in these cases, a spring is used between the bottom of the slot and the inner edge of the vane to bias the vane radially outwardly. J

However, these arrangements impose limitations on the design of the pump as to capacity and outlet pressure, and are limited to low speed, low pressure, fixed capacity pumps. For example, the pump having vanes extending axially through the rotor cannot be designed for variable delivery, so that variation in netoutput can be attained only by recirculating excess pumped fluid from the outlet back to the inlet, with the resultant disadvantages of low overall efiiciency and excessive heating of the pumped fluid.

These considerations are of particular importance in the design of rotary vane pumps for aircraft applications, such as fuel transfer and supply pumps for turbojet engines wherein the fuelpurnped is ,utilized as the lubricant for the pump. As the pumped fuel is not a good lubricant, further limitations are imposed as to allowable bearing loads and relative surface velocities of moving parts.

In my copending application, Serial I-Io. 528,920, filed August 1, 7,-.l9,55, l,have shown, described-and claimeda variable delivery, high lift, positive displacement, con-.

stant pressure rise pump capable of long-timeservicej free operation under adverse conditions, as when used as a submerged fuel transfer pump for aircraft fuel: tanks. "In this pump, centrifugal force is relied upon to;

keep the vane outer edges in contact with the cavity inner surface, and the rotor and cavity, which are both cylin; drical, are mounted for relative axial displacementbe-j tween coaxial and maximum eccentricity positions to vary the stroke of the vanes and thus vary the delivery:-

of the pump. In the pump of my copending application, thevar'ies" are arranged in uniformly angularly spaced slots each.

extending 'in a plane parallel tog-a radial plane of the rotor and each the same distance from such radial. plane;

As compared to a construction in which the vanes are mounted in radial slots, the parallel plane arrangement,

because of the 'angle of the vanes at, their. point'o'f' contact with the pump liner, reduces the load against the latter and thus also the force required to'push the vanes inwardly in their slots in the rotor. This arrangement allows also the use of substant ially longer length. Passage means in the rotor place the inner ends of the slots in communication with the cavity'so that the pressures on the inner and outer edges of the vanes are substantial-ly balanced, and any rem-ainingslight unbalance, depending upon the vane position during rotating of the rotor, supplements or reduces the centrifugal force due to the angular velocity of the rotor. The resultant net force on the vanes urges the vanes outwardly to maintain their outer edges in contact with the cavity surface. This resultant net force is sufiicient to maintain adequate engagement between the vane outer edges and the surface of the pumping cavity'during all normalgoperating conditions of the pump except when the pump is required to operate at very low speeds such as, for example; --at one-sixth of the normal rated design speed. Under extreme operating conditions, such as that mentionedi the centrifugal force may be too low to maintain the vanes in adequate contact with the cavity, surfacejand r-rnustbeaugmented by springs behindth e vanes to raisethe net resultant force on the vanes to a minimum-value sufiicient to insure proper operationat such low, speeds.

Normally, in rotary vane pumps, the rotor and cavity are so interrelated that the vanes-reach their maximum extended position at a point between theinlet'and outlet ports, and are retracted, by engagement ofJheir 'outer edges with the cavity surface, from this point on' toward the outlet port. The vanes reach their maximum retractedposition after passing the'outlet port and at a point as; tween the outlet and inlet ports, being thereafterextended asthey reach ,and pass the inlet port. 1

With such location of the point of maximum extension of the vanes intermediate the inlet and outlet ports, and generally midway therebetween, an analysis of the efie'Q tive side loading pressures against the vanes moving from the inlet port toward the outlet port indicates that, once the trailing vane passes the closing edge of an inlet port and the leading vane has-passed the opening edge of frictional resistancedue to the net. side loadingon the Pa e ted ,Aua- -1 .0.

3 vane resulting from the ditference in inlet and outlet pressures acting on the vane.

In the case of a pump designed for a medium outlet pressure, or of a pump designed for use where the differential between the inlet and outlet pressures is of a medium value, the centrifugal force acting on the vanes at normal operating speeds of the pump is adequately sufficient to maintain the vane outer edges in full contact with the cavity surface as the vanes move from the inlet port toward the point of maximum extension. Once the vane passes the point of maximum extension, it is moved toward the retracted position by the relative convergence of the cavity surface toward the rotor surface, and the force acting to move the vanes inwardly is adequately in excess of the frictional resistance due to the net side loading on the vane.

In my copending application Serial No. 549,606, filed November 29, 1955, now Patent No. 2,885,960, I have shown and described a variable delivery rotary vane pump for high outlet pressures of the order of 1000 p.s.i. or greater. This high pressure pump involves a novel combination of a plurality of pump units, of the type forming the subject matter of my first-mentioned copending application, so arranged that the hydraulic loadings on the bearings for the common shaft of the rotors are reduced to a negligible value, thus providing for long bearing life and maintenance-free operation of the high outlet pressure pump.

However, an analysis of the side loading pressures on the vanes of the high pressure pump indicates that, under certain conditions, the effective centrifugal force acting on the vanes is insuflicient to overbalance the frictional resistance due to the net side loading on the vanes as the latter are moving from the trailing edge of an inlet port toward their position of maximum extension midway between the inlet and outlet ports. Thus, under these conditions, the vanes, during the initial part of their movement from the inlet port to the outlet port will not be extended sufliciently to sealingly engage the cavity surface which, during this part of the vane movement, is at a progressively increasing distance from the surface of the rotor.

One method of correcting this is to reduce the relative eccentricity of the rotor and cavity, and thus the stroke and projected area of each vane. This reduces the side loading pressure on the vanes, but also cuts the pump displacement to a fraction of its previous value necessitating, for equal displacement, an increase in pump size and weight far beyond acceptable limits for its application in aircraft fuel systems.

Another possible solution, usable either alone or with the foregoing impractical solution, is to increase the diameter of the cylinder or cavity inner surface. This would increase the distance between the center of gravity (C.G.) of each vane and the rotor axis, and also increases the weight of each vane. Both increases result in an increase in the centrifugal force acting on the vane. This solution is unacceptable in practice because the rubbing velocity of the vane outer edges against the cavity surface is increased far beyond an acceptable value for long time service free operation of the pump.

In accordance with the present invention, a practical solution is provided by changing the curvature of the surface 'of the cavity, between the closing edge of the inlet port and the point midway between the inlet and outlet ports, where the cavity surface begins to converge toward the rotor surface, so that the cavity surface at the closing edge of the inlet port is at its maximum distance from the rotor surface; this being true in both the fixed and variable displacement embodiments of the pump, irrespective of displacement adjustment of the latter embodiment between maximum and zero displacement. From the closing edge of the inlet port to the point where the cavity surface normally begins to converge toward the rotor surface, the .cavity surface, in accordance with 4 the invention, remains uniformly spaced from the rotor surface or may converge slightly toward the rotor surface.

With the invention arrangement, each vane attains its most extended position at the closing edge of the inlet port and before it is fully subjected to the side loading due to the outlet pressure. From this point toward the outlet port, each vane remains either stationary in its slot or is moved radially inwardly by engagement of its outer edge with the cavity surface.

In a preferred embodiment, the surface of the cavity between the closing edge of the inlet port and a point about midway between the closing edge of the inlet port and the opening edge of the outlet port, where the convergence of the cavity surface toward the rotor surface begins, is formed as a circular are about the axis of the rotor in the case of a fixed displacement pump and, in the case of a variable displacement pump, about the axis of the rotor when the rotor and cavity are in the maximum delivery position. From the starting point of such normal convergence to the point of normal minimum distance between the rotor and cavity surfaces, midway be tween the closing edge of the outlet port and the opening edge of the inlet port, the cavity surface is formed as a circular are about the cavity axis. The remainder of the cavity surface is a transition circular arc joining the first two arcs.

For an understanding of the invention principles, reference is made to the following descriptions of typical embodiments thereof as illustrated in the accompanying drawing. In the drawing:

Fig. l is a plan view of a positive displacement fuel pump embodying the invention;

Fig. 2 is an elevation view of the pump;

Fig. 3 is a sectional view, on the line 33 of Fig. 1, taken through the center of the pumping element;

Fig. 4 is a sectional view, on the line 44 of Fig. 3, taken through the pumping element and the pump displacement adjusting servo-mechanism;

Fig. 5 is an enlarged diametric sectional view through the rotor and cavity of the pump shown in Figs. 1-4, illustrating the relative rotor, cavity and vane positions at maximum displacement, the cavity being an exactly circular cylinder;

Fig. 6 is a graph illustrating the variation in net displacement as a vane travels from the closing edge of the inlet port to the opening edge of the outlet port;

Figs. 7 and 8 are diagrammatic edge views of a vane of the pump of Figs. 1-5 illustrating the forces acting thereon during movement of the vane from the closing edge of the inlet port to a point about midway between the closing edge of the inlet port and the opening edge of the outlet port, at which point the vane has its maximum extension from its rotor slot;

Fig. 9 is a view similar to Fig. 5 illustrating the cavity surface formed in accordance with the invention;

Fig. 10 is a graph similar to Fig. 6 but related to the pump shown in Fig. 9;

Figs. 11 to 15 are views similar to Figs. 7 and 8 but related to a vane of the pump of Fig. 9; and

Fig. 16 is a view similar to Fig. 9 illustrating the invention as incorporated in a six-vane pump.

For illustrative purposes, the invention will be described as incorporated in a variable delivery pump unit of the type forming the subject matter of my copending application Ser. No. 528,920, and which is incorporated in the high pressure variable delivery pump of my copending application Ser. No. 549,606, now Patent No. 2,885,960. Figs. 1 through 4 of the drawings are the same as Figs. 1 through 4 of my copending application Ser. No. 528,920 of which the instant application is a continuation-in-part.

Referring to Figs. 1 through 4, the pump internals are contained within a pair of mating housing members 10 and 20 detachably secured together by bolts 11. Mem- Q. ber 10, which may conveniently be designated the base member, is formed with an aperture'd circular mounting flange .12 by means of which the pump may be mounted at the bottom opening of an aircraft fuel .tank, or on any other suitable support surface. For convenience, housing member 20 will be termed the cap member.

Members and 20 are cooperatively formed to operatively position and support the pumping elements comprising pump rotor 40' and pump body or cylinder 50. For this purpose, members 10 and 20 are formed with facing coaxial stepped cylindrical bores 13 and 23, respectively receiving flanged sleeve bearings 14 and 24. The flange of bearing .24 seats against the bottom of the larger section of bore 23, and the body of bearing 24 has a circumferential groove receiving a packing ring 26 engaging the cylindrical surface of the smaller section of bore 23. A dowel pin 21 seated in cooperating recesses in body member 29 and the flange of bearing 24 restrains this bearing against rotation.

Bearing 14 has a sliding fit in bore 13, with its outer cylindrical surfaces engaging the corresponding surfaces of bore 13. Three springs 27, seated in recesses 17 in base member 10, engage the bottom surface of the flange of the bearing to bias the bearing outwardly of bore 13 to provide a shallow cylindrical space between the inner surface of the flange of the bearing and the bottom of the larger diameter section of the bore;

The springs 27 urge the outer surface of the bearing flange into fluid-tight engagement with a machined face 51 on pump body 50. This forces the pump body toward bearing 24 to maintain a machined face 52 of the pump body in fluidwtight engagement with the outer surface of the flange of bearing 24. The flanges of bearings 14 and 24 thus form the end walls of a cylindrical pumping cavity 55 in body 50, the diameters of the flanges being substantially larger than the diameter of cavity 55. A dowel pin arrangement (not shown), sirnilar to'that including dowel pin 21, restrains bearing 14 against rotation.

Bearings 14 and 24 rot-atably receive the axles or trunnions 41 and 42, respectively, of pump rotor 40. Trunnion 42 has a splined extension 43 extending beyond bearing 24 and slidably engaged in the splined hub 61 of armature or rotor 62 of electric motor 60. A washer 63 and nut 64 retain armature 62 on extension 43, and a thrust washer 65 is disposed between armature 62 and the ends of bearing 24 and o-f-boss 22 surrounding the smaller part of bore 23. The thickness of thrust bearing 65 is such as to allow a slight sliding movement of armature 62 on extension 43. The stator 66 of motor 60 is mounted in a bore 28 in housing member 20, this bore being closed by a cap 25.

When the pump is mounted in the vertical position, thrust bearing 65 normally carries the Weight of the motor armature assembly, but thrust bearing 65 is intended mainly to carry a momentarily increased load due to inertia forces of the mass of the armature assembly resulting from high upward acceleration of modern, fast climbing military aircraft. Thrust loads in the opposite direction, due to fast diving of the aircraft, are taken by the top face of pump rotor 40 against the flange of bearing 24.

One end of pump body 50 has extending therethrough and beyond a hollow cylindrical pivot 53 seating in facing coaxial bores 19 and 29 in members 10 and 20. The bore of pivot 53 communicates with the hollow interior of body 50 through one or more ports 54. Packing rings 56 in circumferential grooves in the trunnions of pivot 53 form seals with the surfaces of bores 19 and 29. The hollow interior of pivot 53 communicates with pump outlet ports 68. While two outlet ports are illustrated, one in member 10 and the other in member 20, the pump may be formed with a single outlet port or one port 68 may be .closed by a suitable plug. The opposite end of pump body 50 has an inlet port 57 to ab t oei cavity 55,;which may,if desired, beextendedby an inlet elbow 67 to'loweri the inlet opening and bring it closer to the bottomof the fuel tank, depending on the pump design. p r

The" inlet port 57 of body 50, or its elbow extension 67, open into a relatively large outwardly opening inlet chamber 70 cooperatively formed between members 10 and 20 and provided with a relatively coarse mesh screen 71 across its open end to prevent any foreign material entering the pump.

The pumping cavity 55 is of larger diameter than pump rotor 40. With body 50 pivoted to oscillate about tubular pivot 53, the axis of pumpcavity 55 can be displaced, with respect to the fixed axis of rotor 40, between a position of maximum eccentricty of the rotor and cavity and a position in which the rotor and its cavity are coaxial or concentric. At the position of maximum eccentricity, the pum'p delivery is a maximum and, at themaxial position, the pump output is zero. '7 Automatic pressure responsive servo-mechanism means are provided, to control the position of cavity 55 relative to rotor 50 to vary the pump output. This means is described more fully hereinafter.

The construction ofpump rotor 40 is best illustrated in Pigs.'4 and'5. Referring to these figures, the body 45 of the rotor has cut therein four longitudinal slots 44 parallel to its-axis and preferably equi-spaced circumferentially of the rotor. Slots 44 may be -in radial planes or in planes parallel to radii. The latter relation is illustrated. The slots 44 are closed at each end by the flanges of bearings 14 and 24, but communicate with pumping cavity 55 through the medium of longitudinal bores or passages 46, at the inner end of each slot, each connected by one or more pressure balancing passages 47 to the exterior surface of the rotor. An axial passage 4S extends through trunnion 41 and into the rotor body 45. This passage is closed at its inner end and is intended only to reduce the weight of the pump rotor.

Each slot 44 slidably mounts a pump vane 75 in the form of a fiat plate, these vanes, when the pump is in operation, being projected by centrifugal force into en gagement with the inner surface of cavity 55. The inner ends of vanes 75 are substantially flat, with rounded edges, but the outer ends of thevanes are bevelled, as at 76, to form a chisel shaped edge or nose having a slightly rounded point 77 at the leading edge of the vane. The chisel shaped nose of each vane has the advantage of keeping the line of contact between the vane and the surface of cavity 55 always in the same location "adjacent the leading edge of the vane, which is not the case when the vane has a rounded nose, as is usually provided. Furthermore, the chisel shaped nose plays an important part in the pressure balancing of the vane.

'In the position of pump body 50 shown in Fig. 4, wherein the rotor 40 and cavity 55 have their maximum concentricity, the surface of rotor 40 nearest the surface of cavity 55 is only a few 'thousandths of an inch therefrom, and the pump delivery is a maximum. If body 50 is swung counterclockwise, about the axis of pivot 53 until cavity 55 is concentric with rotor 40, the pump delivery will be zero as there will no longer be any reciprocation of vanes 75 and thus no change in the volume of the spaces enclosed by rotor 40, cavity 55, vanes 75, and the flanges of bearings 14 and 24. p 7

Pump body 50 is continuously biased in. a. clockwise directionby a button 30 slidably mounted in a socket 3'1 threaded into a wallof member 20 and enclosing a spring 32 urging button 30 against a seat 33 on the pump body. The exact angular positionofbodv 50,,iand thus the pump delivery rate, is controlled automatically by a hydraulic servo-mechanism operating responsive to the pressure rise through thev pump; i.e., the differential between the pump inlet and outlet pressures. This mechanism operates on a ball bearing35 mounted on a pin 34 in forked arms 59 projecting from body 50 substantially diametrically of cavity 55 from seat 33.

Roller bearing 35 engages a sloping cam or wedge surface 78 on a cam 80 operated by a hydraulic actuator 85. This actuator comprises a cylinder 81, a piston 82 fitting closely therein, and a cam' guide sleeve 83. Cylinder 81 seats in a bore 36 in member 20 and has a flange 84 secured by bolts or screws 86 to member 20. Guide sleeve 83 is seated through a bore 37 in member 20, eccentric to bore 36, and has a flange 87 secured bybolts or screws 88 to member 20; An annular groove 89 embracing cylinder 81 is formed partly in the cylinder and partly in the surface of bore 36. Packing rings 91 are disposed on either side of groove 89. Groove 89 com= municates with the space 92, between piston 82 and cylinder head 84, by means of holes 93. Piston 82 has an abutment 94 engaging head 84 to limit movement of the piston toward the head, and carries packing rings 96 engaging cylinder 81.

Cam 80, which is cylindrical in cross section, is centered in bore '97 of sleeve 83 and biased toward piston 82 by a spring 98 engaging a seat 99 on the closed outer end of bore 97. The opposite end of cam 80 carries a hardened button 101 resting against a hardened spacer 102 in piston 82. Shims 103 are positioned between spacer 102 and the piston head.

In order to prevent surface engagement between cam 80 and bore 97 to reduce friction, the bore 97 has a slightly large diameter than that of cam 80. Cam 80 is held in spaced concentric relation with bore 97 by longitudinally spaced ball bearings 90 rotatable on pins 104 in slots 106 in sleeve 83. Bearings 90 engage a groove 107 in cam 80 parallel to the cam axis, this groove forming a track for bearings 90. Bearings 90 project from slots 106 only enough to space cam 80 slightly from the surface of bore 97, and are maintained in contact with groove 107 by bearing 35 pushing against cam 80 about midway between and opposite bearings 90. Contact between cam 80 and bone 97 at right angles to the plane of rollers 35 and 90 is prevented by engagement of rollers 90 with the side walls of groove 107. Accurate centering of the cam 80 in bore 97 is further assured by correct parallelism of pins '34 and 104, and by correct transverse parallelism of wedge surface 78 and the bottom of groove 107.

Shims 103 are used to adjust the longitudinal position of cam 80 to limit the maximum clockwise displacement of body 50 so as to maintain at least a minimum clear ance between rotor 40 and cavity 55 at the position of maximum eccentricity; The limit of outward movement of piston 82 is provided by engagement of the piston with the accurately machined inner end of bore 37. This limits counter-clockwise movement of body 50 to the position in which rotor 40 and cavity 55 are concentric. Operation of actuator 85 is controlled by a pilot valve assembly to maintain the pressure rise through the pump at a preset value, as described in my copending application Ser. No. 528,920.

Referring more particularly to Fig. 5, when rotor 40 is mounted in cavity 50 and the two parts are in' eccentric or output producing relation, there is an eccentric space 72 between the rotor and cavity divided by vanes 75 into four parts S-1, 8-2, -3, and 8-4 of nearly equal circumferential length. In the rotor position illustrated in Fig. 5, the edges 77 of vanes 75-1 and 75-2 contact the wall of cavity 55 exactly at the edges defining cavity inlet port 57, and the edges of vanes 75-3 and 75-4 contact the cavity wall exactly at the edges defining cavity outlet port 58.

In this position of the rotor 40, portion 8-1 of space 72 is at inlet pressure, with the inlet pressure acting directly against the outer end of vane 75-2 and, through passages 46 and 47, against the inner-end of this vane. The pressures on the outer and inner ends of vane 75-2 are'thus balanced, and this balance would be 100% effective if vane edge 77 were actually a sharp edge. However, for praoticalreasons, edge 77 is rounded to a very small radius, thus slightly reducing the area of the outer end of the vane subjected to inlet pressure as indicated by the dimension lines. This slight reduction, resulting in a very slight pressure ditferential on the ends of the vane, has no detrimental effect on the pump operation. However, if the vanes had the usual fully rounded nose, the pressure differential on the inner and outer ends would be substantial and would have a very substantial effect on the pump operation. With such a fully rounded nose, the line of contact moves constantly, during a revolution of the rotor, between a minimum and a maximum distance from the leading edge of the vane.

The other three vanes are likewise pressure balanced. In the illustrated'position of rotor 40, the ends of vane 7 5-3 are both at the pressure existing in part 8-2 of space 72, and the ends of vane 75-1 are at the pressure existing in part S-4 of space 72; as vanes 75-2 and 75-3 shut oii space S-2 and varies 75-4 and 75-1 shut ott space S-4. Both ends of vane 75-4 are at the outlet pressure at outlet port 58, through the medium of passages 46 and 47.

If rotor 40 is turned a few degrees clockwise, vanes 75-1 and 75-3 will clear the edges of inlet port 57 and outlet port 58, respectively, thus placing part 8-4 of space 72 at inlet pressure and part S2 at outlet pres sure. at outlet pressure, as before. The pressures on the inner and outer ends of each of the four vanes remain substantially balanced.

As the pressures on the inner and outer ends of the vanes 75 is balanced for any position of rotor 40, and for any position of the vane edges 77 relative to the ports 57 and 58, there is no pressure differential operating to move the vanes in their slots 44 in either direction, except for the very slight and negligible pressure differential due to the slight rounding of edges 77. Thus, the only practically eifective force acting to move vanes 75 outwardly in their slots is the centrifugal force resulting from rotation of rotor 40 during operation of the pump.

This location of the opening and closing edges of the ports with relation to the points of contact of the vanes with the wall of cavity 55, as just explained, constitutes the ideal theoretical timing of the ports in a fixed displacement vane pump. In practice, in order to allow for machining tolerances in the cutting of the ports through the cavity wall, the chordal length of the ports is made somewhat less to make certain what a vane always seals the gap between pump ports and that there can be no direct communication between inlet and outlet ports. This remark applies to both the fixed and the variable displacement versions of the pump, but in the latter this intentional deviation from the theoretical ideal timing has to be made greater; in the variable displacement pump the change of the eccentricity of the rotor with respect to the cavity, when adjusting the pump displacement, makes it necessary to work out a compromise in the timing of the ports so that in any one of the extreme operating conditions, meaning of maximum and zero output, there is still maintained a seal between the ports.

In order to analyze the pressure and delivery conditions existing during operation of the pump of Figs. 1 through 5, it will be assumed that the vanes 75 and 1'' wide and 1.167 long, and have 0.025" radius at the leading side of the nose vane. Thesteel vanes have a weight of 0.057 lb. and the diameter of the bore of cavity, cylinder, or liner 55 is 3.000. The assumed operating conditions include an inlet pressure of l0 psi. an outlet pressure of 1000 psi, a pump speed of 4000 r.p.rn., and a maximum eccentricity of rotor 40 and bore 55 of 0.109. The positionof the rotor and cavity in Fig. 5 is that of maximum eccentricity and thus of maximum delivery.

From the closing edge 57-2 inlet port 57 to the opening edge 58-1 of outlet port 58, and from the closing Part S-l remains at inlet pressure and part S3 edge58-2 of the outlet portto the opening edge 57-1 of the inlet port, the cavity 55 has an unbroken surface overits whole width. The inlet and. outlet'ports are cut through the cavity or liner surface only from '7-1 to 57-2, and from 58-1 to 58-2, and are preferably at an angle to the axis of the liner. At the ports, the bearing loads on the noses of the vanes are at their lowest value, as will be demonstrated.

With vane 75-2 positioned with its line of contact on its nose just past closing edge 57-2 of the inlet port, and with vane 75-3 just past opening edge 58-1 of the outlet port, space S-2 is open to space 8-3. The outlet pressure of 1000 p.s.i. acts against the'leading side of vane 75-2, and the inlet pressure of p.s.i. acts against the trailing side thereof. In this position, the center of gravity (06.) of vane 75-2 is 1.125" from theaxis of rotation 140 ofrotor 40. The radius through the (3.6. of vane 75-2 forms anangle of 49 46' with the side of the vane.

Referringto Fig. 7, the centrifugal force CFacting on vane 75-2 is equal to W X r X C, where:

W=0.057 lb.

Factor #C=454,656 (at 4000 rpm.)

Centrifugal force CF thus equals 29.15 lbs.

The component CF-1 of centrifugal force acting parallel to the side of vane 75-2 is equal to CF cosine of the angle between the radius through CG. and the side of vane 75-2, or CF-1:29.15 lbs. X 64583=l8.82 lbs.

Against the leading side of vane 75-2, there is a fluid pressure of 0.458 X 1000, or 458 lbs, and against the trailing side, there is a load of 0.391 X 10 or 3.91 lbs, leaving a net load of 454.09 lbs. perpendicular to the side of the vane.

To find the frictional resistance of vane 75-2 against the side of its slot 44:

P2+P1=643.57 lbs. Assuming a coefiicient of friction of 0.15, the force F required to slide the vane in its slot l 5 In the illustrated position of vane 75-2, the inlet pressure of 10 p.s.i., through passage 47, in communication with space S-1, and passage 46, applies a fluid pressure load of 1.87 lbs. against the inner end of vane 75-2. At the outer end of vane 75-2, the rounded nose of the vane provides an area of .0387 sq. in. subjected to the 1000 p.s.i. outlet pressure, and an area of .1488 sq. inas ubend of vane 75-2. This leaves a net load of 40.19-1.87

lbs., or 38.32 lbs. tending to push the vane inwardly.

Consequently, the net force needed to be overcome to move the vane outwardly is 96.53 +38.32 lbs., or 134.85 lbs. As the available centrifugal force on vane 75-2 in the illustrated position is only 18.82 lbs., centrifugal force cannot be relied upon to project vane 75-2 to its maximum'extent at point 115 midway between closing edge 57-2 of the inlet port and opening edge 58-1 of the outlet port.

. At point-115, where vane 75-2'is shown in dotted lines, the vane has reached its maximum projection and thereafter is moved inwardly'by the surface of bore 55 converging toward the surface of rotor 40. The dista'nce of the'center of gravity of vane 75-2 from 140 is now 1.150", and a radius through its C.G. makes an angle of 48 28 43" with the sideof the vane. Vane'75-2,

estate;

at point 115, has its maximum' area mg. Referringto Fig.8. an making the same calculations as before thecentrifugal forceiC.F. is1now 19.75 lbs., the force resisting movement of'the vane in'its slot is 108.0 lbs., and the netflfiuid loading on the zvane ends, tending to pushthe vaneinwardly; is 40.19 lbs., The total force'required to move the vane outwardly is 108+40.19- lbs. or 148.19 lbs.,' substantially in excess of the available centrifugal force-of 19.75 lbs. q

Thus, in the movementof thejvane 75-2 from closing edge 57-2 of the inlet port to point 115, there is not sufficient effective, centrifugal force on the vane .to'inoye the latter outwardly from the intermediate position at. edge 57-2 to the maximum extended position at point115.

l i ed to side 1556 From point around to edge 57-2, there is no probf lem in moving vane 75-2, as will be made clear.

Referring to Figs. 5 and 6, the pump displacement is the' difierence between the volume of space S-2 and the volume of space S-4 foreach of the vanes 75 in making one revolution. To determine the displacement curve, space 8-2 is divided into 18 segments, as shown, by the division lines A to T, and space 8-4 is correspondingly divided. In the latter case the division lines have been omitted so as not to obscure the drawing. Segment AB of space S-2 is the one extending from edge 57-2 of the inlet port, and the corresponding segment of space S-4 extends from edge 58-2 of the outlet port. Pulsations in the delivery will occur four times each revolution for the four-vane pump of Fig. 5.

When the difference in between the volumes of corresponding segments of spaces S2 and 8-4 are plotted as a function of displacement, the curve 116 of Fig. 6 is obtained. In thisfigure, the displacement cross sectional area varies from a maximumof .0283 sq. in. to a minimum of .0198 sq. in., with a mean of .0249 sq. in. The net displacement thus varies from 13.6% above the mean to 20.14% below the mean, which is Well within acceptable limits for intended applications. However, as has been shown, centrifugal force alone is not enough to move the vanes outwardly in their slots while the vanes move from edge 57-2 to point 115.

In order to retain the desirable advantages of using centrifugal force alone to move the vanes outwardly in their slots, the contour of the cavity or liner 55 is modified so that the vanes attain their maximum extension or projection at the closing edge 57-2 of inlet port 57 and, in moving toward point 115 either remain stationary in their slots or are moved slightly inwardly. From point 115 on- Wardly to point 120, midway between edge 58-2 of outlet port 58 and edge 57-1 of inlet port 57, the vanes are moved inwardly in the same manner as in Fig. 5. From point 120 to edge 57-2 of the inlet port, the vanes are moved outwardly at a greater rate than in Fig. 5 so as to attain, at edge 57-2, an extension at least equal to that at point 115 in Fig. 5.

Fig. 9 illustrates a modified surface contour of liner, cylinder or cavity 55 in accordance with the invention. In Fig. 9, the rotor axis is at and theaxis of the major part of the surface of cavity 55 is at 155, corresponding to cylinder axis of Fig. 5. A third center or axis is indicated at 160. v

In Fig. 9, the cavity surface 141 ,from the closing edge 57-2 of inlet port 57 to the normal point 115 of maximum vane projection is a cylindrical surface about rotor with a 3" diameter cavity, the radius of surface 141 is 1.609", the radius of surface 156 is 1.500", and the radius of surface 161 is 1.50 a 1 With this surface contour of cavity 55', the vanes will reach their maximum extension at the trailing edge 572 of inlet port 57, will remainstation-ary until they reach point-115, will be moved inwardly from point 115 to point 120, and will move outwardly from point 120 to edge 57-2. An analysis of the vane loading will illustrate that, in the arrangement of Fig. 9, the centrifugal force acting on the vanes 'is sufficient to move them outwardly in firm contact with the cavity surface as the vanes move from point 120 to edge 57-2. 1

The physicalfactors applicable to the arrangement Fig. are as follows:

Referring to Fig. 9, with a vane 75 just in advance of edge 57-2, space S1 is open to space S-4 and thus at the inlet pressure of p.s.i. The center of gravity of the vane is 1.152 from axis 140, and the radius through C.G. makes, an angle of 48 27 with the side of the vane. Both sides and both the inner and outer ends of the vane are subjected to the inlet pressure, so there is perfect balance of the hydraulic forces on the vane. Using the same method of calculation as in the case of Fig. 5, the component CF1 of centrifugal force acting parallel to the side of the vane is 20.49 lbs., this force being the net forcing urging the vane outwardly to firmly engage surface 161.

When the vane 75-2 moves just past edge 57-2, the preceding vane 753 clears outlet port edge 58-1, so that space S-2 is now at the 1000 p.s.i. outlet pressure. The applicable conditions are illustrated in Fig. 11, with the centrifugal force component CF-l remaining the same as before at 20.49 lbs. Calculating as before, the difference between the hydraulic loadings on the leading and trailing sides of the vane is a net of 422.11 lbs. against the leading side. P-Z equals522.26 lbs. P-1 equals 100.26 lbs., and P-2 plus P-1 equals 622.52 lbs. With the assumed friction coefficient of 0.15, the force F required to slide the vane in its slot is 93.38 lbs.

This value of F is much greater than the value of CF-l, but the vane is already at its maximum extended position. The inlet pressure of 10 p.s.i. exerts a pressure of 1.87 lbs. against the inner end ofthe vane, and the combined effect of the inlet pressure and the 1000 p.s.i. outlet pressure against the outer end of the vane is 96.91 lbs. against the end of the vane resting on surface 141 the vane inwardly.

However, inward movement of the vane is resisted by CF-l equalling 20.49 lbs. and F equalling 93.38 lbs., or a total of 113.87 lbs. This thus leaves a net force of 18.83 lbs. which keeps the vane locked in its uttermost extended position and these conditions hold true as the vane moves from port edge 57-2 to point 115. Exactly at this point, when the vane will start moving inwardly but before the reduction of vane extension has changed loading conditions on the vane, it takes a pressure of 8.83 lbs. against the end of the vane resting on surface 141 to start the vane moving inwardly. 7

As the vane moves from point 115 to 'a position just in advance of opening edge 58-1 of outlet port 58, the surface 156 moves the vane inwardly, thus decreasing the value of CF-l. However, the areas of the vane exposed to hydraulic loading are also decreased, so the net loading of the vane outer edge against surface 156 remains at a favorable value. The conditions just prior to the vane clearing edge 58-1 are as shown in Fig. 12.

Referring to this figure, the distance between the center of gravity of the vane and axis is 1.126", and the radius through C.G. makes an angle of 49 51 with the side of the cane. Calculating as before, CF-l, is 19.47 lbs., the net hydraulic loading against the leading side of the vane is 386.5 lbs., P-2 is 463.30 lbs., P-l is 76.80 lbs., P-2 plus P-l is 540.10 lbs., and F is 81.00 lbs. The total load resisting inward movement of the vane is CF- plus F, or 100.47 lbs. The hydraulic loading against the vane inner end, still at inlet pressure, is 1.87 lbs., and the hydraulic loading against the vane outer end is 94.64, resulting in a net hydraulic force against the outer end of the vane of 92.77 lbs. The net bearing pressure against the surface 156 is thus 100.47 lbs. less 92.77 lbs., or 7.70 lbs.

As the vane moves past edge 581, space S 2 is opened to outlet pressure so that the hydraulic loads on the sides and ends of the vane are balanced, passages 46 and 47 communicating with space S-2. The only force acting on the vane is the component CF-l of the centrifugal force, equal to 19.47 lbs. This is the bearing load of the vane against the ported surface 156.

The hydraulic loads on the vane remain balanced as the vane moves to a position just in advance of outlet port trailing edge 58-2. At this point, the vane center of gravity is 1.018" from axis 140, and the radius through C.G. makes an angle of 57 51" with the side of the vane. Calculating as before, CF-l is now 16.87 lbs. Between edges 58-1 and 58-4, the bearing load of the vane against the ported cavity surface is only the force CF-l, which as shown has decreased from 19.47 lbs. to 16.87 lbs.

The conditions existing when the vane clears edge 53--2 are as depicted in Fig. 13. In this vane position, space S-4 is closed off from the outlet pressure of space S-3 and, as the preceding vane has cleared edge 57-1, space S4 is at the 10 p.s.i. inlet pressure. The pressures against the sides of the vane are reversed, with the outlet pressure against the trailing side and the inlet pressure against the leading side.

The net fluid load against the trailing side of the vane is 163.96 lbs., and can be considered as concentrated against the outer end of the vane as the full bearing length of the vane is engaged in slot 44. The force F resisting movement of the vane is thus 164 0.15 or 24.60 lbs. The net fluid pressure load pushing the vane outwardly is 187.5 lbs., minus .979 lbs. minus 89.6 lbs., or 96.92 lbs. The net bearing load against the cavity surface 156 is thus 16.87+96.9224.60 lbs., or 89.19 lbs.

The conditions existing when the vane reaches the point 120, where the vane has its maximum retraction, are illustrated in Fig. 14. The vane center of gravity is at 0.994" from axis 140, and the-radius through C.G. makes an angle of 59 48 with the side of the vane. Calculating as before, the effective component of centrifugal force CF-l is 13.11 lbs. The net fluid loading against the trailing side of the vane is 136 minus 1.62, or 134.38 lbs., and the force F resisting movement of the vane in its slot is 20.16 lbs. vane outwardly is 187.5 minus 1.01 minus 86.1, or 100.4 lbs. The total bearing load of the vane against the surface of cavity 55' is 13.11 plus 100.4 minus 20.16, or 93.35 lbs.

When the vane reaches a position'just in advance of opening edge 57--1 of the'inlet port, it has moved outwardly somewhat as shown in Fig. 15. The effective radius of the vane center of gravity is now 1.025" and the radius through C.G. makes an angle of 57 33 with the side of the vane. CF-l as calculated is now 14.75 lbs., the net fluid loading on the trailing side of the vane is 166.98 lbs., and the resistance to vane movement in its slot is 25.05 lbs. The net fiuid pressure acting to move the blade outwardly is 102.56 lbs., so that the bearing load of the vane against surface 161 is 92.26 lbs.

With the vane justpast edge 5 7 -1, and the following The net fluid pressure acting to push the vane closing space S4 to the outlet pressure, the hydraulic loading on the vane ends and sidesi's all at the inlet pressure of space S-1, and thus balanced. The only elfective force acting to move the vane outwardly is CF-l which is 14.75 lbs.

While the hydraulic loading 'on the balanced as the vane is moved'from edge 571 to a point just in advance of edge 57-1-2, the vane slides outwardly by about 0.215" so that the centrifugal force, and thus the bearing load against surface 161, increases rapidly to 20.49 lbs.

From the foregoing analysis, it will be apparent that the bearing load of the 'vaneagainst the cavity surface is highest from a point just'beyond port-edge 585-2 to a point just in advance of port edge 57--1: However, as this portion of the cavity surface is unbroken by ports, there is the maximum bearing "surface available to absorb the vane loading. Moving from port edge 58-2 to port edge 57-1, the vane 75 is first pushed inwardly by 0.040" from port edge "58-2 to point 120, and then slides outwardly 0.047 from point 120 to port edge 57-1. This means that the sliding velocity of the vane in its slot is lowest when the bearing load against its outer nose is highest, which is anadvantageous condition. With the circumference of the cavity surface equalling 9.459", the mean rubbing velocity of the vane-against this surface is 3153 feet per minute at 4000 r.p.m. V V

When rotor 40 is revolving,the fluid exerts a wedging action between the rounded nose of the vane and, the cavity surface, which tends to reduce the bearing load of the vane on the cavity surface, reducing the friction and also providing effective lubrication at the line of contact. This wedging action is greatest at the zone of maximum bearing pressure against the cavity surface, between port edges 58-'2"and'57 1'.

f Fig. illustrates thedeliv'ery curve for the arrangement shown in Fig. 9, this curve being calculated in the same manner as the curve of Fig.6. The displacement varies from a maximum value 8.88% above the mean value to a minimum value 19.69% below the mean value. Thus the amplitude of delivery pulsations is even less than with the arrangement shown in Figs. 1-5.

The cavity surface contour of Fig. 9 is not limited to a four-vane rotor, but may be used with a rotor having a greater number of vanes. Fig. 16 illustrates cavity 55 as used with a rotor '40 having six equi-spaced nonradial vanes 75-1 to 75-6. An analysis of the centrifugal forces and pressures involved in Fig. 16 will prove that there is a net force acting to hold the vane against the cavity surface in all positions of the vane around the cavity.

While a different size and weight of vane was assumed in the analysis of Fig. 9, as compared to those of Fig. 5, the same underlying principles apply so that the arrangement of Fig. 9 is equally effective with the vane size and Weight of Fig. 5.

The essence of the invention is in providing a cavity surface contour such that the vanes reach their maxi mum extensions before they are' subjected to the side loading due to the outlet pressure, and thereafter either remain stationary in their slots to point 115 or are moved slightly inwardly. Stated another way, there is no outward movement of the vane required during movement from port edge 57--2 to port edge 581 while the vane is side loaded with the outlet pressure. This results in their being a net force holding the vane against the inward movement while moving to the outlet port, this net force being due in large part to the centrifugal force component parallel to the vane side.

While specific embodiments of the invention have been shown and described in detail to illustrate the application of the invention principles, it will be understood that the invention may be embodied otherwise without departing from such principles.

vane remains What is claimed is:

; 1. 'A rotary vane pump comprising, in combination,

drical surface; a cylindrical pump rotor mounted for ro tation in said cavity with its axis parallel to the axis. of said cavity, the diameter of said rotor being less than that of said cavity; and vanes slidably mounted in circumferentially spaced longitudinal slots in said rotor for engagement of their outer ends with the cylindrical surface of said cavity; the cylindrical surface of said cavity consisting entirely of consecutive circular arcs aboutat least two axes spaced from each other and including the axis of said rotor and the normal axis of-said cavity, being at a maximum distance from the surface of said rotor at the closingj'edge of the inlet port, and converging toward said rotor surface from a point intermediate the closing'edge of the inlet port and the opening edge of the outlet-port to a point intermediate the closing edge of the outlet port and the opening edge of the inlet port; the cylindrical surface of said cavity between the closing edgev of the inlet port and said first mentioned point being a first circular are about the axis of said rotor and, between said points, being a second circular are about the normal axis of saidcavity.

2. A rotary vane pump comprising, in combination, means forming a cylindrical pumping cavity having circumferentially spaced inlet and outlet ports in its cylindrical surface; a cylindrical pump rotor mounted for rotation in said cavity with its axis parallel to the axis of said cavity, the diameter of said rotor being less than that of said cavity;-. and vanes slidably mountedin circumferentially spaced longitudinal slots in said rotor for engagement of th'efi outer ends with the cylindrical surface of said cavity; the cylindrical surface of said cavity consisting entirely of consecutive circular arcs about at least two axes spaced from each other and including the axis of said rotor, the normal axis of said cavity, and a third axis adjacent the cavity axis, being at a uniform maximum distance from the surface of said rotor from the closing edge of the inlet port to a point intermediate the closing edge of the inlet port and the opening edge of the outlet port, converging toward said rotor surface from such point to a point intermediate the closing edge of the outlet port and the opening edge of the inlet port, and diver-ging from the rotor surface from said latter point to the closing edge of the inlet port; the cylindrical surface of said cavity between the closing edge of the inlet and said first mentioned point being a first circular are about the axis of said rotor and, between said points, being a second circular are about the normal axis of said cavity; the cavity surface between said latter point and the closing edge of said inlet port, being a circular transition arc about said third axis and interconnecting said first mentioned circular arcs.

3. A rotary vane pump comprising, in combination, means forming a cylindrical pumping cavity having diametrically opposite inlet and outlet ports in its cylindrical surface; a cylindrical pump rotor mounted for rotation in said cavity with its axis parallel to the axis of said cavity, the diameter of said rotor being less than that of said cavity; and va'nes slidably mounted in circumferentially spaced longitudinal slots in said rotor for engagement of their outer ends with the cylindrical surface of said cavity; the cylindrical surface of said cavity consisting entirely of consecutive circular arcs about at least two axes spaced from each other and including the axis of said rotor and the normal axis of said cavity, being at a maximum distance from the surface of said rotor at the closing edge of the inlet port, and converging toward said rotor surface from a point inter-mediate the closing edge of the inlet port and the opening edge of the outlet port to a point intermediate the closing edge of the outlet port and the opening edge of the inlet port; the cylindrical surface of said cavity between the closing 15 edge of the inlet port and said first mentioned point being a first circular are about the axis of said rotor and, between said points, being a second circular are about the normal axis of said cavity.

4. A rotary vane pump as claimed in claim 3 in which said rotor and cavity are mounted for adjustment of their axes relative to each other to vary the pump displacement; the axis or said first circular are being the axis of the rotor in the position of maximum pump output.

5. A rotary vane pump as claimed in claim 3 in which there are four slots in said rotor; and the circumferential lengths of said ports are equal to the circumferential spacing of the outer ends of adjacent vanes.

6. A rotary vanepump as claimed in claim 3 in which there are six slotsin said rotor.

7. A rotary vane pump comprising, in combination, means forming a cylindrical pumping cavity having diametrically opposite inlet and outlet ports in its cylindrical surface; a cylindrical pump rotor mounted for rotation in said cavity with its axis parallel to the axis of said cavity, the diameter of said rotor being less than that of said cavity; and vanes slidably mounted in circumferentially spaced longitudinal slots in said rotor for engagement of their outer ends with the cylindrical surface of said cavity; the cylindrical surface of said cavity consisting entirely of, consecutive circular arcs about at least two axes spaced from each other and including the axis of said rotor, the normal axis of said cavity, and a third axis adjacent the cavity axis, being at a uniform maximum distance from the surface of said rotor from the closing edge of the inlet port to a point intermediate the closing edge of the inlet port and the opening edge of the outlet port, converging toward said rotor surface from such point to a point intermediate the closing edge of the outlet port and the opening edge of the inlet port, and diverging from the rotor surface from said latter point to the closing edge of the inlet port; the cylindrical surface of said cavity between the closing edge of the' References Cited'in the file of this patent UNITED STATES PATENTS 353,199 Brewer Nov. 23, 1886 963,690 Curtis July 5, 1910 1,200,505 Killman Oct. 10, 1916 1,237,668 Machlet Aug. 21, 1917 1,350,775 Brauer Aug. 24, 1920 1,805,063 Wrona May 12, 1931 1,965,388 Ott July 3, 1934 1,996,875 McCann Apr. 9, 1935 2,045,014 Kenney et al. June 23, 1936 2,057,381 Kenney etal. Oct. 13, 1936 2,073,215 Mann Mar. 9, 1937 2,291,856 Willson Aug. 4, 1942 2,352,941 Curtis July 4, 1944 2,585,406 Reynolds Feb. -12, 1952 2,588,342 Bidwell Mar. 11, 1952 2,599,927 Liver-more June 10, 1952 2,658,456 Wahlmark Nov. 10, 1953 2,708,884 Deschamps May 24, 1955 FOREIGN PATENTS 147,224 Switzerland Aug. 1, 1931 310,980 Germany Feb. 13, 1919 787,204 France July 1, 1935 887,257 France Aug. 9, 1943 

